BACKGROUND OF THE INVENTION
1. Field of the Invention
[0001] The invention relates to a controller and control method, configured to control an
internal combustion engine and a transmission coupled to the output of the internal
combustion engine according to the preamble of claims 1 and 13, respectively and capable
of changing a gear ratio and a shift mode by using a continuously variable transmission
and a gearshift mechanism.
2. Description of the Related Art
[0002] Conventionally, belt-type and traction-type (toroidal-type) transmission mechanisms
are known as a continuously variable transmission mechanism for use in a vehicle transmission.
These continuously variable transmission mechanisms transmit the torque by using the
friction force between a transmitting member such as belt and power roller and a rotor
such as pulley and disc, shearing force of an oil film, and the like, in order to
continuously vary the gear ratio. This restricts the transmittable torque, reduces
the power transmission efficiency when the gear ratio is large or small, and also
limits a practically applicable gear ratio.
[0003] Conventionally, the transmission is not constructed using a continuously variable
transmission mechanism alone but is constructed using combination of a continuously
variable transmission mechanism and a gear mechanism such as a planetary gear mechanism.
One example of such a transmission is described in JP-A-11-504415. The example of
the transmission described in this publication will be described briefly. A driving
pulley of the continuously variable transmission mechanism is coupled to the engine,
and a driven pulley thereof is coupled to a sun gear of the planetary gear mechanism.
A ring gear, which is arranged concentrically with the sun gear, is coupled to the
output shaft of the transmission. A carrier holding a pinion gear meshed with the
sun gear and the ring gear is coupled to the input shaft of the transmission through
the gear mechanism and a clutch. In order to integrally rotate the whole planetary
gear mechanism, a so-called integrating clutch for selectively coupling the prescribed
two rotating elements in the planetary gear, that is, the sun gear and the ring gear,
is provided.
[0004] When the integrating clutch couples the sun gear and the ring gear so as to integrate
the whole planetary gear mechanism, the output torque of the engine is transmitted
from the input shaft of the transmission to the planetary gear mechanism and the output
shaft of the transmission through the continuously variable transmission mechanism.
As a result, the gear ratio of the continuously variable transmission mechanism itself
becomes the overall gear ratio of the transmission. On the other hand, the output
torque of the continuously variable transmission mechanism is transmitted to the sun
gear, and at the same time, the torque is transmitted from the input shaft to the
carrier through the gear mechanism. In this case, the revolutions speed of the sun
gear becomes smaller than that of the carrier as the gear ratio set by the continuously
variable transmission mechanism increases. In other words, the revolution speed of
the ring gear as the output revolution speed becomes larger than that of the carrier.
[0005] The transmission described in the above publication is capable of setting the following
two modes: a mode (shift mode) in which the overall gear ratio of the transmission
increases with increase in the gear ratio of the continuously variable transmission
mechanism; and a mode (shift mode) in which the overall gear ratio of the transmission
decreases with increase in the gear ratio of the continuously variable transmission
mechanism.
[0006] The continuously variable transmission capable of continuously varying the gear ratio
is advantageous in that, when used as a transmission of a vehicle driven by a power
engine, the continuously variable transmission is capable of arbitrarily setting the
revolution speed of the power engine, allowing improvement in power consumption. However,
regarding the technology of cooperatively controlling a power engine and a transmission
capable of changing the shift mode by using a continuously variable transmission mechanism
and a planetary gear mechanism as described in the above publication, problems and
means for solving them have not been sufficiently considered. Moreover, the above
publication does not describe this type of control.
[0007] A generic controller and control method configured to control an internal combustion
engine and a transmission coupled to an output of the internal combustion engine and
including a continuously variable transmission and a gearshift mechanism is known
from DE 196 31 236 A, which shows the features of the preambles of claims 1 and 13.
The controller is capable of selectively setting a first shift mode in which an overall
gear ratio increases with increase in an input/output revolution ratio, that is, a
ratio between an input revolution speed and an output revolution speed of the continuously
variable transmission, and a second shift mode in which the overall gear ratio decreases
with increase in the input/output revolution ratio. The controller comprises an internal
combustion engine control means for obtaining target torque based on control data
including a required output, and controlling load of the internal combustion engine
so as to achieve the target torque.
[0008] A further controller of this type is known from EP-A-0 497 0.38. This reference teaches
to use a second shift mode for a low range, while a first shift mode is used for a
high range.
SUMMARY OF THE INVENTION
[0009] It is an object of the present invention to further develop a controller according
to the preamble of claim 1 and a control method according to the preamble of claim
13 such that fuel consumption and drivability is improved.
[0010] According to the invention, this object is achieved by a controller having the features
of claim 1 and by a control method having the features of claim 13.
[0011] Advantageous further developments are set out in the dependent claims.
[0012] According to the invention, the controller obtains a target input revolution speed
based on the control data and calculates the input/output revolution ratio of the
continuously variable transmission so as to achieve the target input revolution speed.
The controller then determines a shift mode to be set, based on the input/output revolution
ratio calculated by the input/output revolution ratio calculating means
[0013] With this structure, shift shock as well as shift delay can be avoided during switching
between different shift modes that are set in the transmission. Moreover, the engine
can be operated with optimum fuel consumption.
BRIEF DESCRIPTION OF THE DRAWINGS
[0014] Hereinafter, the invention will be described in conjunction with the following drawings.
Each numeral in the figures denotes a corresponding member.
Fig. 1 is a flowchart illustrating an example of the control by a controller according
to the invention;
Fig. 2 is a flowchart illustrating an example of the learning control of the timing
of switching the shift mode;
Fig. 3 is a timing chart illustrating an example of the change of a transmission when
the control of Fig. 1 is conducted;
Fig. 4 is a flowchart illustrating another example of the shift-mode switching control;
Fig. 5 is a timing chart illustrating an example of the change of a transmission when
the control of Fig. 4 is conducted;
Fig. 6 is a skeleton diagram schematically showing an example of a driving system
of a vehicle including a transmission;
Fig. 7 is an alignment chart illustrating shift operation of the transmission;
Fig. 8 is a table showing the correspondence between the running state of a vehicle
and the engaged/released state of each clutch and a brake;
Fig. 9 is a graph showing the relation between the ratio between the input revolution
speed and the output revolution speed of a transmission and the input/output revolution
speed ratio of a continuously variable transmission mechanism; and
Fig. 10 is a block diagram showing an example of a control system for cooperatively
controlling an engine and a transmission based on the required output.
DESCRIPTION OF THE PREFERRED EMBODIMENTS
[0015] Hereinafter, the invention will be described in terms of a specific example. First,
an exemplary driving system (power train) in a vehicle intended to be used in the
invention will be described with reference to Fig. 6. In the illustrated example,
a transmission 2 is coupled to the output of an engine 1, i.e., a power engine, for
example an internal combustion engine. The engine 1 is a power plant for outputting
power by fuel combustion. Accordingly, an engine such as gasoline engine, diesel engine
and natural gas engine can be employed as the engine 1. The engine 1 may be combined
with a motor or a motor generator. The engine 1 is capable of electrically controlling
the load such as throttle opening. The engine 1 including an electronic throttle valve
is exemplarily used herein. The transmission 2 is mainly composed of a belt-type continuously
variable transmission 3 and a single-pinion type planetary gear mechanism 4 serving
as a gearshift mechanism.
[0016] An input shaft 5, an input member, is arranged coaxially with the output shaft of
the engine 1. The input shaft 5 and the engine 1 are coupled to each other through
a damper 6, so that the output shaft of the engine 1 and the input shaft 5 always
rotate together.
[0017] A driving pulley 7, one rotor in the continuously variable transmission 3, is mounted
to the input shaft 5. The driving pulley 7 varies the gap between a fixed sheave and
a movable sheave, that is, the groove width of the driving pulley 7, by moving the
movable sheave with respect to the fixed sheave in the axial direction. Note that
the movable sheave is arranged on the side away from the engine 1, that is, on the
left side in Fig. 6, with respect to the fixed sheave. Accordingly, an actuator 8
for axially moving the movable sheave in both directions is arranged on the back side
of the movable sheave, that is, on the left side in Fig. 6.
[0018] A driven pulley 9, the other rotor in the continuously variable transmission 3, is
arranged on the same plane as the rotation plane of the driving pulley 7. The driven
pulley 9 has the same structure as that of the driving pulley 7. More specifically,
the driven pulley 9 has a fixed sheave and a movable sheave, and varies its groove
width by moving the movable sheave in the forward and backward directions using an
actuator 10. Note that the respective groove widths of the driving pulley 7 and the
driven pulley 9 are controlled so that the groove width of one pulley decreases with
increase in the groove width of the other pulley. The respective groove widths are
varied so as not to change the central position of the driving pulley 7 and the driven
pulley 9 in the axial direction.
[0019] A belt 11, a power-transmission member, is mounted on the driving pulley 7 and the
driven pulley 9. By varying the respective groove widths of the driving pulley 7 and
the driven pulley 9 in the opposite directions, the respective effective diameters
of the driving pulley 7 and the driven pulley 9 on which the belt 11 is mounted vary.
As a result, the input/output revolution ratio γ of the continuously variable transmission
3, that is, the ratio between the revolution speed of the driving pulley 7 (the input
revolution speed of the continuously variable transmission 3) and the revolution speed
of the driven pulley 9 (the output revolution speed of the continuously variable transmission
3) varies continuously. An intermediate shaft 12 is mounted to the driven pulley 9
in order to receive and output the torque to and from the driven pulley 9.
[0020] Hereinafter, the planetary gear mechanism 4 will be described. The planetary gear
mechanism 4 in Fig. 6 includes as rotating elements a sun gear 13, a ring gear 14
and a carrier 15. The sun gear 13 is an external gear. The ring gear 14 is an internal
gear that is arranged concentrically with the sun gear 13. The carrier 15 holds a
pinion gear meshed with the sun gear 13 and the ring gear 14 so that the pinion gear
can rotate on its axis and revolve about the sun gear. The planetary gear mechanism
4 is arranged between the input shaft 5 of the driving pulley 7 and the intermediate
shaft 12 of the driven pulley 9.
[0021] An output shaft 16 extends through the planetary gear mechanism 4 along the central
axis of the driving pulley 7 and the driven pulley 9. The output shaft 16 extends
toward the belt 11 at its one end. This end of the shaft 16 and the ring gear 14 are
coupled together by an appropriate coupling member such as connecting drum. The planetary
gear mechanism 4 includes a direct-coupling clutch Cd for selectively coupling the
ring gear 14 and the sun gear 13. More specifically, the direct-coupling clutch Cd
serves to couple the two rotating elements of the planetary gear mechanism 4, the
ring gear 14 and the sun gear 13, so as to integrally rotate the whole planetary gear
mechanism 4.
[0022] A hollow shaft integrated with the sun gear 13 is rotatably fitted on the outer periphery
of the output shaft 16. The hollow shaft extends in the direction away from the direct-coupling
clutch Cd at its one end. A driving gear 17A is mounted at that end of the hollow
shaft. A driven gear 17B is also mounted at the end of the hollow shaft located away
from the direct-coupling clutch Cd. The driven gear 17B meshes with the driving gear
17A. In other words, the hollow shaft and the intermediate shaft 12 are coupled to
each other through the driving gear 17A and the driven gear 17B. Note that the driving
gear 17A and the driven gear 17B together form a decelerating mechanism from the intermediate
shaft 12 toward the hollow shaft.
[0023] A driving gear 18A is rotatably fitted on the outer periphery of the input shaft
5. The planetary gear mechanism 4 includes a coupling clutch Ch for selectively coupling
the driving gear 18A and the input shaft 5. A driven gear 18B meshed with the driving
gear 18A is rotatably fitted on the outer periphery of the hollow shaft. The driving
gear 18A has a larger diameter than that of the driven gear 18B. The driving gear
18A and the driven gear 18B form a decelerating mechanism from the driving gear 18A
toward the driven gear 18B. More specifically, provided that the gear ratio between
the driving gear 17A and the driven gear 17B is α, the gear ratio between the gear
pair of the driving gear 18A and the driven gear 18B is set to (γ
min × α). Note that γ
min is the minimum value of the input/output revolution ratio γ that is set by the continuously
variable transmission 3.
[0024] Accordingly, when the torque from the engine 1 is transmitted to the sun gear 13
through the continuously variable transmission 3, the driving gear 17A and the driven
gear 17B, and at the same time transmitted to the carrier 17 through the driving gear
18A and the driven gear 18B, the sun gear 13 and the carrier 15 rotate at the same
speed. Accordingly, the whole planetary gear mechanism 4 rotates integrally. This
state is the same as the state in which the direct-coupling clutch Cd is engaged.
[0025] The driven gear 18B is coupled to the carrier 15 in the planetary gear mechanism
4. A brake Br is provided as a fixing means for selectively fixing the driven gear
18B and the carrier 15. In the example of Fig. 6, a friction brake, that is, a wet
type multiple-disc brake, is used as the brake Br. Note that the brake Br needs only
be of a friction type. Therefore, a band brake may be used as the brake Br. The only
requirement for the brake Br is the capability to selectively stop rotation of the
carrier 15. Therefore, the brake Br may be arranged coaxially with the input shaft
5 so as to fix the coupling clutch Ch or the driving gear 18A.
[0026] An output gear 19, an output member, is mounted at the other end of the output shaft
16, that is, the end located away from the belt 11. For example, the output gear 19
is meshed with a ring gear 21 of a front differential 20 so as to output torque to
the front differential 20.
[0027] Note that, by way of example, each of the direct-coupling clutch Cd, the coupling
clutch Ch and the brake Br is herein of a hydraulic type. Accordingly, although not
particularly shown in the figures, a hydraulic controller is provided in order to
control these engaging/releasing mechanisms. The driving system includes an electronic
control unit for the transmission (T-ECU) 22 for controlling the engaged/released
state of the engaging/releasing mechanisms and controlling the input/output revolution
ratio γ that is set in the continuously variable transmission 3. The T-ECU 22 is a
unit based on a microcomputer. Detection signals such as vehicle speed, accelerator
opening, oil temperature, input/output revolution speeds of the transmission 2 and
revolution speeds of the driving pulley 7 and the driven pulley 9 are input to the
T-ECU 22. The T-ECU 22 conducts shift-mode switching and shift control as described
below according to the above input signals and pre-stored data and programs.
[0028] The driving system further includes an electronic control unit for the engine (E-ECU)
23 for electrically controlling the load of the engine 1. Like the T-ECU 22, the E-ECU
23 is a unit based on a microcomputer. For example, the E-ECU 23 obtains the target
torque based on the control data including the required output represented by the
accelerator opening and the like, and the running state such as vehicle speed, and
sets the load of the engine 1 so that the output torque reaches the target torque.
The T-ECU 22 and the E-ECU 23 are connected so that they are capable of data communications
with each other.
[0029] The transmission 2 including the continuously variable transmission 3 and the planetary
gear mechanism 4 is capable of conducting shifting operation in two shift modes. One
of the shift modes is a shift mode in which only the shifting function of the continuously
variable transmission 3 is used to set the gear ratio, that is, a shift mode in which
the gear ratio Γ of the transmission 2 increases/decreases with increase/decrease
in the input/output revolution ratio γ of the continuously variable transmission 3.
This shift mode is herein referred to as direct mode or L mode. The other shift mode
is a shift mode in which both the shifting function of the continuously variable transmission
3 and the shifting function of the planetary gear mechanism 4 are used to set the
gear ratio, that is, a shift mode in which the gear ratio Γ of the transmission 2
varies in the direction opposite to that of the input/output revolution ratio γ of
the continuously variable transmission 3. This shift mode is herein referred to as
a power recirculation mode or H mode. Hereinafter, the above two shift modes will
be described in conjunction with the shifting function of the transmission 2.
[0030] First, when starting the engine 1, the direct-coupling clutch Cd, the coupling clutch
Ch and the brake Br are rendered in a released state (i.e., disengaged state). In
the case where the engine 1 drives a hydraulic pump (not shown) for operating the
direct-coupling clutch Cd and the coupling clutch Ch, the above engaging elements
are rendered in the released state without particular control before starting the
engine 1. However, in the case where the driving system includes a means for accumulating
the hydraulic pressure of the hydraulic pump or in the case where another power source
drives the hydraulic pump, the above elements are rendered in the released state by
discharging the hydraulic pressure from the hydraulic pump serving as a means for
engaging the above elements. Accordingly, when starting the engine 1, the carrier
15 will neither function as a reacting element nor an input element since the input
shaft 5 is disengaged from the driving gear 18A and the brake Br is released. Moreover,
the direct-coupling clutch Cd is released and thus the planetary gear mechanism 4
is not integrated. Therefore, no torque will be transmitted from the engine 1 to the
output shaft 16. In other words, the engine 1 is started with the transmission 2 being
rendered in the neutral state.
[0031] Hereinafter, description will be given for the case where the vehicle is started.
When the vehicle is started in the forward direction, the gear ratio Γ of the transmission
2 must be increased as soon as the engine 1 is started. Accordingly, the groove width
of the driving pulley 7 in the continuously variable transmission 3 is maximized so
as to minimize the effective diameter of the driving pulley 7 on which the belt 11
is mounted, and at the same time, the groove width of the driven pulley 9 is minimized
so as to maximize the effective diameter of the driven pulley 9 on which the belt
11 is mounted. Thus varying the respective effective diameters of the belt 11 on the
driving pulley 7 and the driven pulley 9 maximizes the input/output revolution ratio
γ of the continuously variable transmission 3 (γ max). At the maximum input/output
revolution ratio γ max, the sun gear 13 and the ring gear 14 are gradually engaged
with each other by the direct-coupling clutch Cd. In other words, by gradually increasing
the engaging hydraulic pressure for operating the direct-coupling clutch Cd, the sun
gear 13 and the ring gear 14 change from the disengaged state to the completely engaged
state through the slip state. The above operation gradually increases the torque transmitted
from the engine 1, that is, the torque capacity. Therefore, the torque capacity on
the output shaft 16 varies gently, allowing smooth starting of the vehicle.
[0032] Fig. 7 is an alignment chart showing the above state for the planetary gear mechanism
4. With the direct-coupling clutch Cd engaging the sun gear 13 and the ring gear 14,
the whole planetary gear mechanism 4 rotates integrally. Accordingly, the torque is
transmitted from the engine 1 (Eng) to the sun gear 13 through the continuously variable
transmission 3 (CVT). The ring gear 14 serving as an output element and the output
shaft 16 coupled thereto thus rotate in the same direction at the same speed together
with the sun gear 13 serving as an input element. This operating state is shown by
line A in Fig. 7.
[0033] The input/output revolution ratio γ of the continuously variable transmission 3 is
reduced from the above state. More specifically, the groove width of the driving pulley
7 is gradually reduced to increase the effective diameter thereof, and the groove
width of the driven pulley 9 is gradually increased to reduce the effective diameter
thereof. As a result, the relative input revolution speed to the planetary gear mechanism
4 increases gradually. Since the whole planetary gear mechanism 4 rotates integrally,
the revolution speed of the output shaft 16 relative to the engine speed increases
with a change in the input/output revolution ratio γ of the continuously variable
transmission 3. In other words, provided that the vehicle speed does not vary, the
engine speed decreases with decrease in the gear ratio of the continuously variable
transmission 3. Such change in the operating state is shown by line A in Fig. 7 translated
upward, i.e., in the direction in which the revolution speed increases. The state
in which the planetary gear mechanism 4 is rendered in a so-called direct-coupled
state so as to minimize the input/output revolution ratio γ of the continuously variable
transmission 3, i.e., to set the input/output revolution ratio γ to the minimum value
γ
min corresponding to the highest speed, is shown by line B in Fig. 7.
[0034] As described above, the direct mode (L mode) is the state in which the direct-coupling
clutch Cd engages the sun gear 13 with the ring gear 14 and the coupling clutch Ch
disengages the driving gear 18A from the input shaft 5. In the direct mode, a change
in the input / output revolution ratio γ of the continuously variable transmission
3 appears as a change in the overall gear ratio Γ of the transmission 2.
[0035] When the input/output revolution ratio γ of the continuously variable transmission
3 has the minimum value γ
min, the gear ratio between the driving gear 18A and the driven gear 18B is (γ
min × α), where α is the gear ratio between the driving gear 17A and the driven gear
17B, which are between the intermediate shaft 12 and the hollow shaft. Therefore,
the revolution speed of the driving gear 18A matches the engine speed. This allows
engaging of the coupling clutch Ch as well as release of the direct-coupling clutch
Cd without causing any rotational fluctuation in each rotating member and without
changing the output shaft torque from the transmission 2. The minimum value γ
min of the input/output revolution ratio γ is a predetermined input/output revolution
ratio suitable for the shift mode of the invention.
[0036] Such engaging and releasing of the clutches is conducted so that the carrier 15 has
a revolution speed according to the engine speed, and the revolution speed of the
sun gear 13 is varied by the continuously variable transmission 3. This enables setting
of a so-called overdrive state. The overdrive state herein refers to the state in
which the gear ratio Γ of the transmission 2 is set to a value smaller than the minimum
value that can be set by the continuously variable transmission 3 alone.
[0037] The above state is shown by Line C in Fig. 7. In this state, the revolution speed
of the carrier 15 according to the engine speed is maintained and the input/output
revolution ratio γ of the continuously variable transmission 3 is increased so that
the revolution speed of the sun gear 13 is reduced. As a result, the respective revolution
speeds of the ring gear 14 serving as an output element and the output shaft 16 coupled
thereto are increased. In other words, the overall gear ratio Γ of the transmission
2 is further reduced. Provided that the vehicle speed does not change, the engine
speed is reduced. This is called the power recirculation state.
[0038] As described above, the power recirculation mode (H mode) is the state in which the
direct-coupling clutch Cd disengages the sun gear 13 from the ring gear 14 and the
coupling clutch Ch engages the driving gear 18A with the input shaft 5. In this mode,
the overall gear ratio Γ of the transmission 2 varies in the direction opposite to
that of the input/output revolution ratio γ of the continuously variable transmission
3. More specifically, by increasing the input/output revolution ratio γ of the continuously
variable transmission 3, the gear ratio Γ can be set to a value smaller than the value
that can be set by the continuously variable transmission 3 alone.
[0039] Note that, as described above, provided that the input/output revolution ratio γ
of the continuously variable transmission 3 is set to the minimum value γ
min, the whole transmission 2 rotates integrally even when the direct-coupling clutch
Cd is released. This state, which corresponds to the highest speed in the direct mode
(L mode) and the lowest speed in the power recirculation mode (H mode), is common
to each shift mode. In other words, the minimum value γ
min of the input/output revolution ratio γ corresponds to the shift point, that is, the
switch point, from one shift mode to the other mode. Note that the shift point is
determined by the respective gear ratios between the driving gear 17A and the driven
gear 17B and between the driving gear 18A and the driven gear 18B.
[0040] Releasing both the direct-coupling clutch Cd and the coupling clutch Ch, and engaging
the brake Br enable reverse running of the vehicle. More specifically, by engaging
the brake Br, the carrier 15 is fixed in the planetary gear mechanism 4. In this state,
the torque is input to the sun gear 13 through the continuously variable transmission
3. Accordingly, the ring gear 14 serves as an output element, so that the output shaft
16 coupled thereto rotates in the direction opposite to that of the sun gear 13. This
state is shown by line D in Fig. 7.
[0041] Fig. 8 collectively shows the engaged/released state of each engaging/releasing mechanism
for setting the direct mode (L mode), the power recirculation mode (H mode) and the
reverse-running mode. In Fig. 8, the term "range" refers to a running mode of the
vehicle selected by manual operation, where "R" indicates the range for reverse running,
"P" indicates the range for maintaining the stopped state, "N" indicates the range
for setting the neutral state, and "D" indicates the range for forward running. In
Fig. 8, each blank indicates a released state, and "O" indicates an engaged state.
For example, the torque capacity transmitted in the engaged state can be arbitrarily
set by adjusting by an electromagnetic value (not shown) the hydraulic pressure for
operating a clutch to be engaged.
[0042] Fig. 9 shows the relation between the speed ratio set by the transmission 2, that
is, the ratio between the input revolution speed Ni and the output revolution speed
No (No/Ni: an inverse number of the gear ratio Γ) and the input/output revolution
ratio γ of the continuously variable transmission 3. Based on the running state such
as vehicle speed and accelerator opening, the transmission 2 conducts shifting operation
so as to achieve the engine speed of the optimum operating point corresponding to
the minimum fuel consumption, while satisfying the required output. The required output
is represented by a signal indicating accelerator opening, a signal from a cruise
control system, and the like. The accelerator opening refers to the depressing amount
of an accelerator pedal. The cruise control system is set based on the preset vehicle
speed, the distance to a vehicle ahead, and the like.
[0043] The shifting operation includes the following types: shifting operation within each
shift mode as shown by line G1 in Fig. 9; shifting operation across the shift point
as shown by line G2 in Fig. 9, in which the input/output revolution ratio γ before
and after the shifting operation is approximated to the minimum value γ
min; shifting operation across the shift point as shown by line G3 in Fig. 9, in which
the input/output revolution ratio γ before or after the shifting operation is approximated
to the minimum value γ
min; and shifting operation across the shift point as shown by line G4 in Fig. 9, in
which the input/output revolution ratio γ before and after the shifting operation
significantly deviates from the minimum value γ
min.
[0044] Of the above types of shifting operation, the shifting operation shown by line G1
does not involve switching of the shift mode, that is, does not involve switching
of the engaged/released state of the direct-coupling clutch Cd and the coupling clutch
Ch. This shifting operation is conducted by changing only the input/output revolution
ratio γ of the continuously variable transmission 3. Accordingly, the gear ratio of
the transmission 2 varies continuously, and will not vary in a stepwise manner. In
contrast, in the other shifting operations, the engaged/released state of the direct-coupling
clutch Cd and the coupling clutch Ch is switched in order to switch the shift mode.
Therefore, the gear ratio may vary in a stepwise manner. In view of this, the controller
according to one aspect of the invention conducts the shifting operation involving
switching of the shift mode as described below.
[0045] Shifting operation of the transmission 2 is basically controlled so as to achieve
the required output in a more fuel-efficient manner. More specifically, the gear ratio
is controlled so as to achieve the most fuel-efficient engine speed. In this case,
the load of the engine 1 is controlled to achieve the torque satisfying the required
output. In other words, the engine 1 and the transmission 2 are cooperatively controlled
based on the required output. Fig. 10 is a block diagram showing an example of such
control. The target driving force F is obtained based on the accelerator opening Acc
and the vehicle speed V (Block B1).
[0046] The accelerator opening Acc is the control data resulting from electrically processing
the depressing amount of an accelerator pedal (not shown). The accelerator opening
Acc is used as a parameter indicating the request for acceleration or deceleration,
that is, the required output. Accordingly, a signal indicating driving request for
cruise control that maintains a constant vehicle speed may be used instead of the
accelerator opening Acc. The same applies to the vehicle speed. More specifically,
the revolution speed of another appropriate rotating member that corresponds to the
vehicle speed V on a one-to-one basis may be used instead of the vehicle speed V.
[0047] Determination of the target driving force F based on the accelerator opening Acc
and the vehicle speed V is made based on a predetermined map. More specifically, the
relation between the vehicle speed V and the driving force F is predetermined as a
map by using the accelerator opening Acc as a parameter. In this case, the driving
force F is determined so as to reflect the characteristics of a vehicle of interest.
The target driving force F is thus obtained based on the map.
[0048] The target output P is obtained based on the target driving force F thus obtained
in Block B1 and a current vehicle speed V (Block B2). The target output P is the product
of the target driving force F and the vehicle speed V.
[0049] The target engine speed Net corresponding to the target output P is obtained in order
to control the gear ratio Γ (Block B3). As described before, in the steady running
state, the engine speed is controlled according to the optimum operating line (which
is a set of optimum operating points), that is, the optimum fuel consumption line.
Accordingly, the operating state at the target output P is located on the optimum
operating line. In other words, at the target output P, the engine 1 is controlled
to the state based on the optimum fuel consumption line. Therefore, the target engine
speed Net is obtained by using a target engine speed table (graph) defining the output
and the revolution speed based on the optimum fuel consumption line.
[0050] The gear ratio Γ to be set in the transmission 2 is calculated based on the target
engine speed Net and the vehicle speed V or the revolution speed of the output shaft
(Block B4). The input/output revolution ratio γ of the continuously variable transmission
3 and the shift mode are then obtained in order to set the gear ratio Γ (Block B5).
More specifically, if the gear ratio Γ to be set in the transmission 2 is larger than
the minimum value that can be set by the continuously variable transmission 3 alone,
it is determined that L mode should be set. In contrast, if the gear ratio Γ to be
set in the transmission 2 is smaller than the minimum value that can be set by the
continuously variable transmission 3 alone, it is determined that H mode should be
set.
[0051] A shift control means controls the direct-coupling clutch Cd and the coupling clutch
Ch to set the shift mode to L mode or H mode. The input/output revolution ratio γ
of the continuously variable transmission 3 is set to the target value, that is, so
that the engine speed is located on the optimum fuel consumption line (Block B6).
Specifically, the shift control means is the T-ECU 22 in Fig. 6.
[0052] In order to control the engine 1, the target engine torque To is obtained based on
the target output P and a current engine speed Ne (Block B7). For example, the target
engine torque To is obtained by dividing the target output P by the current engine
speed Ne. Note that the equation shown in Fig. 10 is an equation resulting from the
processing in view of the respective units of the values. Accordingly, the angular
velocity of the output shaft of the engine 1 may be used instead of the engine speed
Ne.
[0053] An engine torque control means controls the engine 1 so as to achieve the target
engine torque To thus obtained (Block B8). More specifically, the T-ECU 22 of Fig.
6 controls the fuel injection amount or the opening of an electronic throttle valve.
[0054] As described above, the controller according to one aspect of the invention controls
the gear ratio Γ of the transmission 2, that is, the shift mode and the input/output
revolution ratio γ of the continuously variable transmission 3, based on the required
output. At the same time, the controller controls the load of the engine 1. Such shift
control in so-called cooperative control of the engine 1 and the transmission 2 will
now be described in more detail.
[0055] Fig. 1 is a flowchart illustrating the overall flow of the control. First, the vehicle
speed V and the accelerator opening Acc indicating the required output are detected
(Step S1). The target output shaft torque To is obtained based on the vehicle speed
V and the accelerator opening Acc (Step S2). This corresponds to the control in Block
B1 of Fig. 10.
[0056] The operating point of the engine 1 (E/G), that is, the target engine speed and the
target torque, is determined based on the vehicle speed V and the accelerator opening
Acc (Step S3). The gear ratio Γ of the transmission 2 (TM) is determined based on
the target engine speed (Step S4). The shift mode and the input/output revolution
ratio γ of the continuously variable transmission 3 are determined based on the gear
ratio Γ (Step S5). The control of Steps S3 to S5 corresponds to the control in Blocks
B3, B4 and B5 in Fig. 10.
[0057] Thereafter, the load of the engine 1 is controlled so that output torque Te reaches
the target torque (Step S6). More specifically, the throttle opening is controlled.
[0058] Regarding the control of the transmission 2, whether the shift mode is to be switched
or not is determined (Step S7). More specifically, if the shift mode determined in
Step S5 is different from the actual shift mode, it is determined that the shift mode
is to be switched. In contrast, if the shift mode determined in Step S5 is the same
as the actual shift mode, it is determined that the shift mode is not to be switched.
[0059] If NO in Step S7, that is, if it is determined that the shift mode need not be switched
in order to achieve the gear ratio Γ determined in Step S4, only the input/output
revolution ratio γ of the continuously variable transmission 3 is varied so as to
achieve the gear ratio Γ (Step S8). This control is conducted in the same manner whether
the actual shift mode is L mode or H mode.
[0060] If YES in Step S7, that is, if it is determined that the shift mode must be switched
in order to achieve the gear ratio Γ determined in Step S4, the input/output revolution
ratio γ of the continuously variable transmission 3 is varied toward the value suitable
for switching the shift mode, that is, toward the minimum value γ
min (Step S9). For example, when the shift mode is to be switched from L mode to H mode,
working fluid is supplied to the actuator 8 of the driving pulley 7 in Fig. 6 so as
to reduce the groove width of the driving pulley 7 and thus increase the groove width
of the driven pulley 9.
[0061] The input/output revolution ratio γ of the continuously variable transmission 3 is
obtained by detecting the respective revolution speeds of the driving pulley 7 and
the driven pulley 9 and calculating the ratio therebetween. Whether the input/output
revolution ratio γ becomes close to the minimum value γ
min or not is then determined (Step S10). As described before, in the transmission 2
of Fig. 6, it is desirable to switch the shift mode when the input/output revolution
ratio γ of the continuously variable transmission 3 becomes close to the minimum value
γ
min. This is because switching the shift mode by changing the engaged/released state
of the direct-coupling clutch Cd and the coupling clutch Ch at the minimum input/output
revolution ratio γ
min of the continuously variable transmission 3 will not cause any rotational fluctuation,
change in output shaft torque and the like. If NO in Step S10, the routine returns
without any particular control. If YES in Step S10, mode switching control (sequence)
is started (Step S11). The mode switching control will be described later.
[0062] Thereafter, whether the switch sequence is completed or not is determined (Step S12).
If YES in Step S12, the timing of starting the shift-mode switching control is corrected
by learning (Step S13). The learning correction will be described later.
[0063] Note that, although not shown in Fig. 1, after the shift mode is switched as described
above, the input/output revolution ratio γ of the continuously variable transmission
3 is gradually increased from the minimum value γ
min to the value determined in Step S5. Operation of varying the gear ratio Γ, i.e.,
shifting operation, that involves switching of the shift mode is thus completed.
[0064] Fig. 2 is a flowchart specifically illustrating the control in Steps S10 to S13.
After the input/output revolution ratio γ of the continuously variable transmission
3 is varied toward the minimum value γ
min, whether the input/output revolution ratio γ becomes close to the minimum value γ
min or not is determined. This determination is made by the operation of Fig. 2 (Step
S101). More specifically, it is determined that whether or not the product of the
difference between the actual input/output revolution ratio γ
n and the minimum value γ
min (γ
n- γ
min) and the monitoring interval T of the input/output revolution ratio γ
n (i.e., the time interval between the previous monitoring and the current monitoring),
(γ
n - γ
min) · T, divided by a variation in the input/output revolution ratio γ during the time
interval T, (γ
n-1 - γ
n), is equal to or smaller than a prescribed value (Tst - ΔT).
[0065] Tst is the stroke time of a clutch that is engaged to switch the shift mode. In other
words, Tst is the time from when supply of the hydraulic pressure for operating the
clutch is started until pack clearance is eliminated and the torque is actually transmitted,
that is, until the clutch starts having the torque capacity. This time corresponds
to the substantial time to switch the shift mode. This value can be obtained in advance
by experimentation or the like. ΔT is a value that is corrected by learning, and may
either be a positive value or a negative value.
[0066] Consequently, in Step S101, the time for the input/output revolution ratio γ of the
continuously variable transmission 3 to reach the minimum value γ
min is estimated, and whether or not the estimated time is equal to or less than the
time required to engage the engaging clutch is determined. If NO in Step S101, the
input/output revolution ratio γ of the continuously variable transmission 3 is not
close to the minimum value γ
min. Therefore, the routine returns without particular control. If YES in Step S101,
the switch sequence is started (Step S11).
[0067] The switch sequence includes the following two controls. One of the controls is control
of the standby state. In the case where the direct-coupling clutch Cd and the couple
clutch Ch is a hydraulic multiple-disc clutch, the standby state is the state right
before engaging the clutch, in which the hydraulic pressure for engaging is increased
and the torque starts having the torque capacity. The other control is control from
the time when the clutch actually starts engaging after the control of the standby
state until the clutch is completely engaged. For example, each control in the switching
sequence may be conducted based on a timer. Accordingly, determination that the sequence
is completed (Step S12) can be made when the counting of the timer is completed.
[0068] As shown in Step S102 of Fig. 2, the switch sequence is started when the estimated
time for the input/output revolution speed γ of the continuously variable transmission
3 to reach the minimum value γ
min matches the stroke time of the hydraulic piston in the engaging clutch minus ΔT.
Accordingly, the switch sequence is completed after the input/output revolution ratio
γ of the continuously variable transmission 3 reaches the minimum value γ
min. In other words, in the process of switching the shift mode, the input/output revolution
ratio γ of the continuously variable transmission 3 is held at the minimum value γ
min for a prescribed time.
[0069] The γ
min duration is then detected. The γ
min duration is a time period from when the input/output revolution ratio γ calculated
based on the respective revolution speeds of the driving pulley 7 and the driven pulley
9 obtained by a revolution speed sensor reaches the minimum value γ until the engaging
clutch starts having the torque capacity after the so-called standby state. Whether
or not the γ
min duration is equal to or less than a predetermined lower limit TL is then determined
(Step S131). If NO in Step S131, whether or not the γ
min duration is equal to or larger than a predetermined upper limit TU is determined
(Step S132).
[0070] If NO in Step S132, the γ
min duration is within a prescribed range defined by the lower limit TL and the upper
limit TU. Therefore, it is determined that the timing of substantially starting the
shift-mode switching control is proper. In this case, the learning correction value
ΔT is not changed.
[0071] If YES in Step S131, the timing when the engaging clutch substantially starts engaging
is too early. In this case, the learning correction value ΔT is changed to the sum
of the learning correction value ΔT and a predetermined value β (Step S133). This
means that the reference value for determination in Step S101 (Tst - ΔT) is reduced.
Therefore, the shift-mode switch sequence will not be started until the input/output
revolution ratio γ of the continuously variable transmission 3 becomes very close
to the minimum value γ. This prevents the shift mode from being switched before the
input/output revolution ratio γ of the continuously variable transmission 3 reaches
the value suitable for switching the shift mode, and thus prevents shocks from being
generated.
[0072] If YES in Step S132, the timing when the engaging clutch substantially starts engaging
is late. In this case, the learning correction value ΔT is changed to the learning
correction value ΔT minus a predetermined value β (Step S134). This means that the
reference value for determination in Step S101 (Tst - ΔT) is increased. Therefore,
the shift-mode switch sequence will be started before the input/output revolution
ratio γ of the continuously variable transmission 3 becomes very close to the minimum
value γ. This prevents the input/output revolution ratio γ of the continuously variable
transmission 3 from being held at the minimum value γ
min suitable for switching the shift mode for an excessively long time, and also prevents
the driver from having a feeling of shift delay. Alternatively, this prevents the
operating point of the engine 1 from significantly deviating from the optimum fuel
consumption line, and also prevents generation of rotational fluctuation upon restoring
the input revolution speed to the target value and thus generation of inertial torque.
Therefore, shocks can be prevented from being generated due to such rotational fluctuation
and inertial torque.
[0073] Fig. 3 exemplarily shows a change in input revolution speed Nin, output revolution
speed No, input/output revolution ratio γ, gear ratio (Nin/No = Γ), clutch hydraulic
pressure for operating a clutch, and output shaft torque To. In Fig. 3, it is assumed
that the vehicle speed is increased with constant accelerator opening and the above
control is conducted during the shifting operation that involves switching of the
shift mode. The output revolution speed No of the transmission 2 gradually increases
with increase in vehicle speed V. The gear ratio Γ is thus gradually reduced in order
to hold the operating point of the engine 1 on the optimum fuel consumption line.
Since the shift mode is L mode, the gear ratio Γ is reduced by reducing the input/output
revolution ratio γ of the continuously variable transmission 3.
[0074] In this process, the target shift mode is determined, and the continuously variable
transmission 3 is controlled toward the input/output revolution ratio γ
min suitable for shifting the shift mode. Provided that the condition of Step 101 in
Fig. 2 is satisfied at time t1 in Fig. 3, the switch sequence is started. First, a
hydraulic pressure for engaging of the engaging clutch (in this case, the coupling
clutch Ch that is engaged in H mode) starts being increased. Then, the engaging clutch
is rendered in the standby state near the end of a predetermined time period T2. Note
that the standby state is the state right before the coupling clutch is engaged with
its pack clearance being eliminated. Therefore, by further increasing the engaging
hydraulic pressure, the clutch would start having the substantial torque capacity.
At time t2 in the standby state, the input/output revolution ratio γ of the continuously
variable transmission 3 reaches the minimum value γ
min, and the standby state is maintained.
[0075] The standby period T2 ends at time t3. At time t3, the hydraulic pressure for engaging
the engaging clutch, i.e., the coupling clutch Ch, is further increased, so that the
coupling clutch Ch starts being engaged. In addition, the engaging hydraulic pressure
is discharged from the releasing clutch (in this case, the direct-coupling clutch
Cd). At time t4, that is, after a predetermined time period T1 from time t3, the engaging
clutch, that is, the coupling clutch Ch, is completely engaged, and the shift mode
is switched to H mode. The time period T1 from when the coupling clutch Ch substantially
starts having the torque capacity until it is completely engaged is set so as to suppress
or prevent problems such as pulsation due to abrupt change in hydraulic pressure.
[0076] After time t4 when the shift mode is substantially switched to H mode, the input/output
revolution ratio γ of the continuously variable transmission 3 is gradually increased
so that the engine speed reaches the target value. At time t6, the input revolution
speed Nin reaches the target value. Shifting operation involving switching of the
shift mode is thus completed.
[0077] Accordingly, in the above shifting process, the input/output revolution ratio γ of
the continuously variable transmission 3 is held at the minimum value γ
min during the period from time t2 to time t4. The vehicle speed and thus the output
revolution speed No increase during this period as well. Therefore, the engine speed,
that is, the input revolution speed Nin, increases away from the target value (solid
line), as shown by the chain line in Fig. 3. This means that the operating point of
the engine 1 deviates from the optimum fuel efficient line.
[0078] However, the controller of the invention corrects the timing of starting the switch
sequence by learning so that the γ
min duration, i.e., the period from time t2 to time t3, is within the range defined by
the lower limit TL and the upper limit TU. As a result, the γ
min duration is reduced as much as possible. This prevents the input revolution speed
Nin from significantly deviating from the target value, that is, prevents the operating
point of the engine 1 from significantly deviating from the optimum fuel consumption
line. Therefore, degradation in fuel consumption can be prevented or suppressed.
[0079] As shown by the chain line in Fig. 3, the output shaft torque To varies only slightly.
Therefore, shock generation and degradation in drivability can be prevented or suppressed.
Moreover, only slight rotational fluctuation occurs when the actual input revolution
speed Nin is controlled to the target value. Therefore, inertial force and shocks
or vibration resulting from the inertial force is prevented or suppressed. In this
respect as well, degradation in drivability can be prevented.
[0080] Note that, in the above control, the shift-mode switching control is conducted at
the minimum input/output revolution ratio γ
min. Therefore, variation in output shaft torque To resulting from simultaneously switching
the engaged/released state of the direct-coupling clutch Cd and the coupling clutch
Ch as well as shock generation resulting from such variation are prevented.
[0081] The change shown by the dashed line in Fig. 3 results from starting switching of
the shift mode after confirming that the input/output revolution ratio γ reached the
minimum value γ
min suitable for switching the shift mode. In such control, the minimum input/output
revolution ratio γ is held at the minimum value γ
min until time t5 that is later than time t4. This increases the γ
min duration and thus increases deviation of the input revolution speed Nin from the
target value. As a result, the operating point of the engine 1 significantly deviates
from the optimum fuel consumption line, degrading the fuel consumption. Moreover,
rotational fluctuation resulting from controlling the input revolution speed Nin to
the target value and thus the inertial torque resulting from such fluctuation are
increased, whereby shocks may possibly be generated.
[0082] The control shown in Figs. 2 and 3 is effective when the hydraulic pressure of the
direct-coupling clutch Cd and the coupling clutch Ch can be directly controlled by,
e.g., a linear solenoid valve (not shown) and the like in an accurate manner. This
control can be similarly applied to a transmission that adjusts a hydraulic pressure
for engaging the direct-coupling clutch Cd and the coupling clutch Ch or a hydraulic
pressure for releasing them by means of an orifice and an accumulator (both of which
are not shown). In other words, the timing of starting supply and discharge of the
hydraulic pressures for engaging and releasing the direct-coupling clutch Cd and the
coupling clutch Ch is controlled based on an instruction given by the T-ECU 22 during
the shift-mode switch sequence. It is therefore effective to correct the timing of
starting supply and discharge of the hydraulic pressures for engaging and releasing
the direct-coupling clutch Cd and the coupling clutch Ch, based on the γ
min duration, that is, a period set by subtracting preset operating time of the accumulator
from the preset maximum time to complete the switching operation. In the former case,
control of Fig. 4 is also possible.
[0083] The control example of Fig. 4 is conducted instead of Steps S10 to S12 of Fig. 1.
This control does not include learning control. First, while the input/output revolution
ratio γ of the continuously variable transmission 3 is being changed toward the minimum
value γ
min suitable for switching the shift mode, the time for the input/output revolution ratio
γ to reach the minimum value γ
min is estimated. This estimation can be conducted in the same manner as that of Step
S101 in Fig. 2. Whether or not the estimated time is equal to or less than the time
Tst corresponding to the piston stroke time of the engaging clutch is determined (Step
S111).
[0084] If NO in Step S111, the routine returns without particular control. If YES in Step
S111, a stroke instruction for the engaging clutch (in the illustrated example, the
coupling clutch Ch that is engaged in H mode) is output (Step S112). According to
this instruction, the hydraulic pressure is supplied to the coupling clutch Ch so
as to move the piston forward against the mounting load of a return spring (not shown)
and the like, and to hold the coupling clutch Ch in the state right before it starts
having substantial torque capacity (standby state). This control is conducted by increasing
the hydraulic pressure to a predetermined value required for engaging the clutch.
Note that the hydraulic pressure is directly controlled by a linear solenoid value
and the like.
[0085] In parallel with the above process, the input/output revolution ratio γ is detected
(monitored). After the input/output revolution ratio γ reaches the minimum value γ
min, whether or not the γ
min duration is equal to or larger than a predetermined upper limit TU is determined
(Step S113). In order to prevent erroneous determination, it is preferable to determine
whether the input/output revolution ratio γ reaches the minimum value γ
min or not by reading the value twice, e.g., detecting a signal from the revolution speed
sensor twice. The upper limit TU can be set as the minimum time period required to
confirm the determination that the input/output revolution ratio γ reached the minimum
value γ
min.
[0086] If NO in Step S113, the routine returns without particular control. If YES in Step
S113, an instruction to increase the hydraulic pressure for engaging the engaging
clutch, that is, the coupling clutch Ch, is output (Step S114). Moreover, an instruction
to reduce the hydraulic pressure for engaging the releasing clutch (in the illustrated
example, the direct-coupling clutch Cd) is output (Step S115). In other words, in
order to switch the shift mode from L mode to H mode, the engaging hydraulic pressure
is increased so as to substantially engaging the coupling clutch Ch, and at the same
time, the hydraulic pressure is reduced so as to release the direct-coupling clutch
Cd. These hydraulic pressures are directly controlled by, e.g., a linear solenoid
value and the like. The tendency (gradient) of change in these hydraulic pressures
is set so that the hydraulic pressures change as soon as possible within the range
that does not cause pulsation due to abrupt change in hydraulic pressure. For example,
the gradient may be approximated to that in the case where the hydraulic pressures
are controlled by mean of the orifice and the accumulator as described in the control
example of Figs. 2 and 3.
[0087] Fig. 5 shows an example of the control of Fig. 4. In this example, the shift mode
is switched from L mode to H mode, and at the same time, the gear ratio is varied
when the vehicle speed increases with constant acceleration opening. The chain line
in Fig. 5 indicates a change resulting from the above control. The solid line in Fig.
5 indicates the target value or an ideal change. The dashed line in Fig. 5 indicates
a change in the control other than the control according to one aspect of the invention.
As described above, in the control according to one aspect of the invention, the timing
when the input/output revolution ratio γ reaches the minimum value γ
min within the piston stroke time Tst is detected, and a hydraulic pressure for engaging
the engaging clutch starts being increased. In this state, it is confirmed that the
input/output revolution ratio γ reached the minimum value γ
min· At time t3 when the γ
min duration reaches the upper limit TU, the hydraulic pressure for engaging the engaging
clutch is increased, and the hydraulic pressure is discharged from the releasing clutch.
[0088] Accordingly, the shift-mode switching by changing the engaged/released state of the
direct-coupling clutch Cd and the coupling clutch Ch is conducted at the minimum input/output
revolution ratio γ
min of the continuously variable transmission 3. This prevents the gear ratio from changing
soon after switching the engaged/released state of each clutch like so-called clutch-to-clutch
shift operation. Moreover, the time period from time t2 when the input/output revolution
ratio γ of the continuously variable transmission 3 reaches the minimum value γ
min suitable for switching the shift mode to time t3 when the shift mode is substantially
switched, that is, the γ
min duration, can be reduced. This reduces deviation of the input revolution speed Nin
from the target value and thus reduces deviation of the operating point of the engine
from the optimum fuel consumption line. As a result, degradation in fuel consumption
can be prevented or suppressed, and also shift shock as well as shift delay can be
avoided.
[0089] Thus, the power engine and the transmission including the continuously variable transmission
coupled to the output of the power engine are controlled so as to satisfy the required
output and to reduce the fuel consumption. As a result, drivability and fuel consumption
can be improved. Moreover, the overall gear ratio of the transmission varies as continuously
as possible, so that the intrinsic characteristics of the continuously variable transmission
can be effectively obtained. Therefore, drivability and fuel consumption can be improved.
[0090] The correspondence between the above specific example and the invention will now
be described briefly. The functional means of Step S6 in Fig. 1 and the functional
means of Block B8 in Fig. 10 correspond to the power engine control means according
to one aspect of the invention. The functional means of Step S5 in Fig. 1 and the
functional means of Block B5 in Fig. 10 correspond to the input/output revolution
ratio calculating means (S5) according to one aspect of the invention. The functional
means of Step S5 in Fig. 1 and the functional means of Block B5 in Fig. 10 correspond
to the shift-mode determining means according to one aspect of the invention.
[0091] Note that the transmission capable of setting two transmission modes, L mode and
H mode, is described in the above specific example. However, the invention is not
limited to such a specific example. And the internal combustion engine is adopted
as a power plant in the above specific example. However, the invention is not limited
to such a specific example. The invention is also applicable to a controller for a
transmission capable of setting three or more shift modes. The only requirement for
the transmission applied to the invention is the capability to set a plurality of
shift modes by a continuously variable transmission and a gearshift mechanism, and
the invention is not limited to the structure of the transmission of Fig. 6.
1. A controller configured to control an internal combustion engine (1) and a transmission
(3, 4) coupled to an output of the internal combustion engine (1) and including a
continuously variable transmission (3) and a gearshift mechanism (4), and the controller
is capable of selectively setting a first shift mode (L) in which an overall gear
ratio (Γ) increases with increase in an input/output revolution ratio (γ), that is,
a ratio between an input revolution speed and an output revolution speed of the continuously
variable transmission (3), and a second shift mode (H) in which the overall gear ratio
(Γ) decreases with increase in the input/output revolution ratio (γ), comprising:
an internal combustion engine control means (6) for obtaining target torque based
on control data including a required output, and controlling load of the internal
combustion engine (1) so as to achieve the target torque;
characterized by
an input/output revolution ratio calculating means (S5) for obtaining a target input
revolution speed based on the control data and on an optimum fuel consumption line
and for calculating the input/output revolution ratio of the continuously variable
transmission (3) so as to achieve the target input revolution speed; and
a shift-mode determining means (S5) for determining a shift mode to be set from the
first shift mode (L) and the second shift mode (H) based on the input/output revolution
ratio (γ) calculated by the input/output revolution ratio calculating means (S5),
wherein said second shift mode (H) is set if said overall gear ratio (Γ) is smaller
than a minimum possible input/output revolution ration (γ) of only said continuously
variable transmission (3).
2. A controller according to claim 1,
characterized in that
an input shaft of the continuously variable transmission (3) is coupled to an output
shaft of the internal combustion engine (1), the gearshift mechanism (4) is a planetary
gear mechanism, a sun gear (13) of the planetary gear mechanism is engaged so as to
rotate in a direction opposite to that of an output shaft of the continuously variable
transmission (3), a carrier (15) of the planetary gear mechanism is capable of being
selectively engaged or released so as to rotate in a direction opposite to that of
the output shaft of the internal combustion engine (1), the sun gear (13) and a ring
gear (14) are capable of being selectively integrated or disengaged, the first shift
mode is set by releasing the carrier (15) and the output shaft of the internal combustion
engine (1) and integrating the sun gear (13) and the ring gear (14), and the second
shift mode that defines a range of the gear ratio that is different from that of the
first shift mode is set by disengaging the sun gear (13) and the ring gear (14) and
engaging the carrier (15) and the output shaft of the internal combustion engine (1).
3. A controller according to claim 1 or 2,
characterized in that
the shift mode is switched when the input/output revolution ratio calculated by the
input/output revolution ratio calculating means (S5) is within the range of the overall
gear ratio defined by a shift mode other than a current shift mode.
4. A controller according to any one of claims 1 through 3,
characterized by
further comprising a continuously variable transmission control means (S12) for switching
the shift mode and setting the calculated input/output revolution ratio (γ) after
varying a current input/output revolution ratio to a predetermined value suitable
for switching the shift mode.
5. A controller according to claim 4,
characterized in that
the continuously variable transmission control means (S12) controls the overall gear
ratio (Γ) so as to achieve a most efficient revolution speed of the internal combustion
engine (1), and the internal combustion engine (1) control means controls the internal
combustion engine (1) so that the internal combustion engine (1) generates torque
according to the required output.
6. A controller according to claim 4,
characterized by further comprising:
an estimating means (S101, S111) for estimating time for the input/output revolution
ratio (γ) to reach the predetermined value when varying the input/output revolution
ratio (γ) while involving switching of the shift mode; and
a shift-mode switching control means for determining timing of starting shift-mode
switching control based on the time estimated by the estimating means (S101, S111).
7. A controller according to claim 6,
characterized in that
the estimating means (S101, S111) estimates the time for the input/output revolution
ratio (γ) to reach the predetermined value based on a variation per unit time in the
input/output revolution ratio (γ).
8. A controller according to claim 7,
characterized in that
the shift mode switching control means integrates or disengages the sun gear (13)
and the ring gear (14), and engages or releases the carrier (15) and the output shaft
of the internal combustion engine (1) based on the time estimated by the estimating
means.
9. The controller as claimed in either of claims 6 and 8,
characterized in that
the controller further comprises a revolution sensor for detecting an input revolution
speed and an output revolution speed of the continuously variable transmission (3);
and
a timer for measuring a time interval for detecting the revolution speeds, and when
the following equation is satisfied, it is determined that the switching control is
to be started:
where T is the time interval,
γ
n-1 and γ
n are the input/output revolution ratio calculated from the revolution speeds detected
at the time interval T, γ
min is the predetermined input/output revolution ratio, and Tst is time required to switch
the shift mode.
10. A controller according to claim 4, characterized by further comprising a shift-mode switching control means for changing the timing of
substantially starting the shift-mode switching control based on duration of the predetermined
input/output revolution ratio in a process of varying the input/output revolution
ratio while involving switching of the shift mode.
11. A controller according to claim 10,
characterized in that
the duration of the predetermined input/output revolution ratio is corrected based
on past operation record of the duration.
12. A controller as claimed in either of claims 10 and 11,
characterized in that
the controller further comprises a revolution sensor for measuring an input revolution
speed and an output revolution speed of the continuously variable transmission (3);
and
a timer for measuring a time interval for detecting the revolution speeds, and when
the following equation is satisfied, it is determined that the switching control is
to be started:
where T is the time interval,
γ
n-1 and γ
n are the input/output revolution ratio calculated from the revolution speeds detected
at the time interval T, γ
min is the predetermined input/output revolution ratio,
Tst is time required to switch the shift mode, and
ΔT is a correction value based on the past operation record of the duration.
13. A method for controlling an internal combustion engine (1) and a transmission (3,
4), wherein said transmission coupled to an output of the internal combustion engine
(1) includes the continuously variable transmission (3) and a gearshift mechanism
(4), the method being capable of selectively setting a first shift mode (L) in which
an overall gear ratio (Γ) increases with increase in an input/output revolution ratio
(γ), that is, a ratio between an input revolution speed and an output revolution speed
of the continuously variable transmission (3), and a second shift mode (H) in which
the overall gear ratio (Γ) decreases with increase in the input/output revolution
ratio (γ), comprising the step of:
obtaining target torque based on control data including a required output, and controlling
load of the internal combustion engine (1) so as to achieve the target torque;
characterized by the steps of:
obtaining a target input revolution speed based on the control data and on an optimum
fuel consumption line and calculating the input/output revolution ratio(γ) of the
continuously variable transmission (3) so as to achieve the target input revolution
speed; and
determining a shift mode to be set from the first shift mode (L) and the second shift
mode (H) based on the calculated input/output revolution ratio, wherein said second
shift mode (H) is set if said overall gear ratio (Γ) is smaller than a minimum possible
input/output revolution ration (γ) of only said continuously variable transmission
(3).
14. A control method according to claim 13,
characterized in that
an input shaft of the continuously variable transmission (3) is coupled to an output
shaft of the internal combustion engine (1), the gearshift mechanism (4) is a planetary
gear mechanism, a sun gear (13) of the planetary gear mechanism is engaged so as to
rotate in a direction opposite to that of an output shaft of the continuously variable
transmission (3), a carrier (15) of the planetary gear mechanism is capable of being
selectively engaged or released so as to rotate in a direction opposite to that of
the output shaft of the internal combustion engine (1), the sun gear (13) and a ring
gear (14) are capable of being selectively integrated or disengaged, the first shift
mode is set by releasing the carrier (15) and the output shaft of the internal combustion
engine (1) and integrating the sun gear (13) and the ring gear (14), and the second
shift mode that defines a range of the gear ratio that is different from that of the
first shift mode is set by disengaging the sun gear (13) and the ring gear (14) and
engaging the carrier (15) and the output shaft of the internal combustion engine (1).
15. A control method according to claim 13 and 14,
characterized in that
the shift mode is switched when the input/output revolution ratio calculated by the
input/output revolution ratio calculating means (S5) is within the range of the overall
gear ratio defined by a shift mode other than a current shift mode.
16. A control method according to any one of claims 13 through 15,
characterized in that
the shift mode is switched, and the calculated input/output revolution ratio is set
after a current input/output revolution ratio is varied to predetermined value suitable
for switching the shift mode.
17. A control method according to claim 17,
characterized in that
the overall gear ratio is controlled so as to achieve a most efficient revolution
speed of the internal combustion engine (1), and the internal combustion engine (1)
is controlled so that the internal combustion engine (1) generates torque according
to the required output.
18. A control method according to claim 16,
characterized by further comprising the steps of:
estimating time for the input/output revolution ratio (γ) to reach the predetermined
value when varying the input/output revolution ratio (γ) while involving switching
of the shift mode; and
determining timing of starting shift-mode switching control based on the estimated
time.
19. A control method according to claim 18,
characterized in that
the time for the input/output revolution ratio (γ) to reach the predetermined value
is estimated based on a variation per unit time in the input/output revolution ratio
(γ).
20. A control method according to claim 19,
characterized in that
the shift mode is switched by integrating or disengaging the sun gear (13) and the
ring gear (14), and engaging or releasing the carrier (15) and the output shaft of
the internal combustion engine (1) based on the time estimated by the estimating means.
21. The control method as claimed in either of claims 18 and 20,
characterized in that
the controller further comprises a revolution sensor for measuring an input revolution
speed and an output revolution speed of the continuously variable transmission (3);
and
a timer for measuring a time interval for detecting the revolution speeds, and when
the following equation is satisfied, it is determined that the switching control is
to be started:
where T is the time interval,
γ
n-1 and γ
n are the input/output revolution ratio calculated from the revolution speeds detected
at the time interval T, γ
min is the predetermined input/output revolution ratio,
Tst is time required to switch the shift mode.
22. A control method according to claim 16,
characterized by further comprising the step of changing the timing of substantially starting the
shift-mode switching control based on duration of the predetermined input/output revolution
ratio in a process of varying the input/output revolution ratio while involving switching
of the shift mode.
23. A control method according to claim 22,
characterized in that
the duration of the predetermined input/output revolution ratio is corrected based
on past operation record of the duration.
24. A control method as claimed in either of claims 22 and 23,
characterized in that
the controller further comprises a revolution sensor for measuring an input revolution
speed and an output revolution speed of the continuously variable transmission (3);
and
a timer for measuring a time interval for detecting the revolution speeds, and when
the following equation is satisfied, it is determined that the switching control is
to be started:
where T is the time interval,
γ
n-1 and γ
n are the input/output revolution ratio calculated from the revolution speeds detected
at the time interval T, γ
min is the predetermined input/output revolution ratio,
Tst is time required to switch the shift mode, and
ΔT is a correction value based on the past operation record of the duration.
1. Steuerung, die konfiguriert ist, um einen Verbrennungsmotor und ein Getriebe (3, 4)
zu steuern, das mit einem Ausgang des Verbrennungsmotors (1) gekoppelt ist und ein
kontinuierlich variables Getriebe (3) und einen Gangschaltmechanismus (4) hat, wobei
die Steuerung fähig ist, wahlweise einen ersten Schaltmodus (L), in dem ein Gesamtgetriebeverhältnis
(Γ) mit einer Erhöhung eines Eingangs-/Ausgangsumdrehungsverhältnisses (γ) steigt,
das heißt ein Verhältnis zwischen einer Eingangsdrehzahl und einer Ausgangsdrehzahl
des kontinuierlich variablen Getriebes (3), und einen zweiten Schaltmodus (H) zu setzen,
in dem das Gesamtgetriebeverhältnis (Γ) mit einer Erhöhung des Eingangs-/Ausgangsumdrehungsverhältnisses
(γ) sinkt, mit:
einer Verbrennungsmotorsteuereinrichtung (6) zum Erhalten eines Solldrehmoments auf
der Grundlage von Steuerdaten einschließlich einem erforderlichen Ausgang und zum
Steuern einer Last des Verbrennungsmotors (1), um das Sollmoment zu erhalten;
gekennzeichnet durch
eine Eingangs-/Ausgangsumdrehungsverhältnisberechnungseinrichtung (S5) zum Erhalten
einer Solleingangsdrehzahl auf der Grundlage der Steuerdaten und einer optimalen Kraftstoffverbrauchslinie
und zum Berechnen des Eingangs-/Ausgangsumdrehungsverhältnisses des kontinuierlich
variablen Getriebes (3), um die Solleingangsdrehzahl zu erreichen; und
einer Schaltmodusbestimmungseinrichtung (S5) zum Bestimmen eines Schaltmodus, der
von dem ersten Schaltmodus (L) und dem zweiten Schaltmodus (H) auf der Grundlage des
Eingangs-/Ausgangsumdrehungsverhältnisses (γ) zu setzen ist, das
durch die Eingangs-/Ausgangsumdrehungsverhältnisberechnungseinrichtung (S5) berechnet wird,
wobei der zweite Schaltmodus (H) gesetzt wird, wenn das Gesamtgetriebeverhältnis (Γ)
kleiner als ein minimal mögliches Eingangs-/Ausgangsumdrehungsverhältnis (γ) von nur
dem kontinuierlich variablen Getriebe (3) ist.
2. Steuerung nach Anspruch 1,
dadurch gekennzeichnet, dass
eine Eingangswelle des kontinuierlich variablen Getriebes (3) mit einer Ausgangswelle
des Verbrennungsmotors (1) gekoppelt ist, der Getriebeschaltmechanismus (4) ein Planetengetriebemechanismus
ist, ein Sonnenrad (13) des Planetengetriebemechanismus in Eingriff ist, um in eine
Richtung entgegengesetzt zu der einer Ausgangswelle des kontinuierlichen variablen
Getriebes (3) zu drehen, ein Träger (15) des Planetengetriebemechanismus fähig ist,
wahlweise in Eingriff zu sein oder von dem Eingriff gelöst zu sein, um in eine Richtung
entgegengesetzt zu der der Ausgangswelle des Verbrennungsmotors (1) zu drehen, das
Sonnenrad (13) und ein Hohlrad (14) fähig sind, wahlweise eingegliedert oder außer
Eingriff zu sein, der erste Schaltmodus durch Lösen des Trägers (15) und der Ausgangswelle
des Verbrennungsmotors (1) und Eingliedern des Sonnenrads (13) und des Hohlrads (14)
gesetzt wird und der zweite Schaltmodus, der einen Bereich des Getriebeverhältnisses,
der von dem des ersten Schaltmodus verschiedenen ist, definiert, durch außer Eingriff
Setzen des Sonnenrads (13) und des Hohlrads (14) und Eingreifens des Trägers (15)
und der Ausgangswelle des Verbrennungsmotors (1) gesetzt wird.
3. Steuerung nach Anspruch 1 oder 2,
dadurch gekennzeichnet, dass
der Schaltmodus umgeschaltet wird, wenn das Eingangs-/Ausgangsumdrehungsverhältnis,
das durch die Eingangs-/Ausgangsumdrehungsverhältnisberechnungseinrichtung (S5) berechnet
wird, innerhalb des Gesamtgetriebeverhältnisses ist, das durch einen anderen Schaltmodus
als einen gegenwärtigen Schaltmodus definiert ist.
4. Steuerung nach einem der Ansprüche 1 bis 3, ferner
gekennzeichnet durch
eine Steuereinrichtung eines kontinuierlich variablen Getriebes (S12) zum Umschalten
des Schaltmodus und Setzen des berechneten Eingangs-/Ausgangsumdrehungsverhältnisses
(γ) nach Variieren eines gegenwärtigen Eingangs-/Ausgangsumdrehungsverhältnisses auf
einen vorgegebenen Wert, der zum Umschalten des Schaltmodus geeignet ist.
5. Steuerung nach Anspruch 4,
dadurch gekennzeichnet, dass
die Steuereinrichtung des kontinuierlich variablen Getriebes (S12) das Gesamtgetriebeverhältnis
(Γ) steuert, um eine effizienteste Drehzahl des Verbrennungsmotors (1) zu erreichen
und die Verbrennungsmotorsteuereinrichtung den Verbrennungsmotors (1) steuert, so
dass der Verbrennungsmotor ein Drehmoment gemäß dem erforderlichen Ausgang generiert.
6. Steuerung nach Anspruch 4, ferner
gekennzeichnet durch,
eine Abschätzeinrichtung (S101, S111) zum Abschätzen einer Zeit für das Eingangs-/Ausgangsumdrehungsverhältnis
(γ), um den vorgegebenen Wert zu erreichen, wenn das Eingangs-/Ausgangsumdrehungsverhältnis
(γ) variiert wird, während ein Umschalten des Schaltmodus involviert ist; und
eine Schaltmodusumschaltsteuereinrichtung zum Bestimmen einer Zeitgebung eines Startens
einer Schaltmodusumschaltsteuerung auf der Grundlage der Zeit, die durch die Abschätzeinrichtung (S101, S111) abgeschätzt wird.
7. Steuerung nach Anspruch 6,
dadurch gekennzeichnet, dass
die Abschätzeinrichtung (S101, S111) die Zeit für das Eingangs-/Ausgangsumdrehungsverhältnis
(γ) abschätzt, um den vorgegebenen Wert auf der Grundlage einer Veränderung je Zeiteinheit
in dem Eingangs-/Ausgangsumdrehungsverhältnis (γ) zu erreichen.
8. Steuerung nach Anspruch 7,
dadurch gekennzeichnet, dass
die Schaltmodusumschaltsteuereinrichtung das Sonnenrad (13) und das Hohlrad (14) eingliedert
oder den Eingriff voneinander löst, und den Träger (15) und die Ausgangswelle des
Verbrennungsmotors (1) auf der Grundlage der Zeit, die durch die Abschätzseinrichtung
abgeschätzt wird, in Eingriff bringt oder von dem Eingriff löst.
9. Steuerung nach entweder Anspruch 6 oder 8,
dadurch gekennzeichnet, dass
die Steuerung ferner einen Drehzahlsensor zum Erfassen einer Eingangsdrehzahl und
einer Ausgangsdrehzahl des kontinuierlichen variablen Getriebes (3); und
einen Zeitmesser zum Messen eines Zeitintervalls zum Erfassen der Drehzahlen hat,
und wenn die nachstehende Gleichung erfüllt ist, bestimmt wird, dass die Umschaltsteuerung
zu starten ist:
wobei T das Zeitintervall ist, γ
n-1 und γ
n das Eingangs-/Ausgangsumdrehungsverhältnis sind, das aus den Drehzahlen berechnet
wird, die bei dem Zeitintervall T erfasst werden, γ
min das vorgegebene Eingangs-/Ausgangsumdrehungsverhältnis ist, und Tst eine Zeit ist,
die erforderlich ist, um den Schaltmodus umzuschalten.
10. Steuerung nach Anspruch 4, ferner
gekennzeichnet durch eine
Schaltmodusumschaltsteuereinrichtung zum Ändern der Zeitgebung im Wesentlichen eines
Startens der Schaltmodusumschaltsteuerung auf der Grundlage einer Dauer des vorgegebenen
Eingangs-/Ausgangsumdrehungsverhältnisses in einem Prozess eines Variierens des Eingangs-/Ausgangsumdrehungsverhältnisses,
während ein Umschalten des Schaltmodus involviert ist.
11. Steuerung nach Anspruch 10,
dadurch gekennzeichnet, dass
die Dauer des vorgegebenen Eingangs-/Ausgangsumdrehungsverhältnisses auf der Grundlage
einer vergangenen Betriebsaufzeichnung der Dauer korrigiert wird.
12. Steuerung nach entweder Anspruch 10 oder 11,
dadurch gekennzeichnet, dass
die Steuerung ferner einen Drehzahlsensor zum Messen einer Eingangsdrehzahl und einer
Ausgangsdrehzahl des kontinuierlich variablen Getriebes (3); und
einen Zeitmesser zum Messen eines Zeitintervalls zum Erfassen der Drehzahlen hat,
und wenn die nachstehende Gleichung erfüllt ist, es bestimmt wird, dass die Umschaltsteuerung
zu starten ist:
wobei T ein Zeitintervall ist, γ
n-1 und γ
n das Eingangs-/Ausgangsdrehverhältnis sind, das aus den Drehzahlen berechnet wird,
die bei dem Zeitintervall T erfasst werden, γ
min das vorgegebene Eingangs-/Ausgangsumdrehungsverhältnis ist, Tst eine Zeit ist, die
erforderlich ist, um den Schaltmodus umzuschalten, und ΔT ein Korrekturwert auf der
Grundlage der vergangenen Betriebsaufzeichnung der Dauer ist.
13. Verfahren zum Steuern eines Verbrennungsmotors (1) und eines Getriebes (3, 4), wobei
das Getriebe, das mit einem Ausgang des Verbrennungsmotors (1) gekoppelt ist, das
kontinuierlich variable Getriebe (3) und einen Getriebeschaltmechanismus (4) hat,
wobei das Verfahren fähig ist, wahlweise einen ersten Schaltmodus (L), in dem ein
Gesamtgetriebeverhältnis (Γ) mit einer Erhöhung eines Eingangs-/Ausgangsumdrehungsverhältnisses
(γ) steigt, das heißt ein Verhältnis zwischen einer Eingangsdrehzahl und einer Ausgangsdrehzahl
des kontinuierlichen variablen Getiebes (3), und einen zweiten Schaltmodus (H) zu
setzen, in dem das Gesamtgetriebeverhältnis (Γ) mit einer Erhöhung des Eingangs-/Ausgangsumdrehungsverhältnisses
(γ) sinkt, mit dem Schritt:
Erhalten eines Solldrehmoments auf der Grundlage von Steuerdaten einschließlich einem
erforderlichen Ausgang und Steuern einer Last des Verbrennungsmotors (1), um das Solldrehmoment
zu erhalten;
gekennzeichnet durch die Schritte:
Erhalten einer Solleingangsdrehzahl auf der Grundlage der Steuerdaten und einer optimalen
Kraftstoffverbrauchslinie und Berechnen des Eingangs-/Ausgangsumdrehungsverhältnisses
(γ) des kontinuierlich variablen Getriebes (3), um die Solleingangsdrehzahl zu erhalten;
und
Bestimmen eines Schaltmodus, der zu setzen ist, aus dem ersten Schaltmodus (L) und
dem zweiten Schaltmodus (H) auf der Grundlage des berechneten Eingangs-/Ausgangsumdrehungsverhältnisses,
wobei der zweite Schaltmodus (H) gesetzt ist, wenn das Gesamtgetriebeverhältnis (Γ)
kleiner als ein minimal mögliches Eingangs-/Ausgangsgetriebeverhältnis (γ) von nur
dem kontinuierlich variablen Getriebes (2) ist.
14. Steuerverfahren nach Anspruch 13,
dadurch gekennzeichnet, dass
eine Eingangswelle des kontinuierlich variablen Getriebes (3) mit einer Ausgangswelle
des Verbrennungsmotors (1) gekoppelt ist, der Getriebeschaltmechanismus (4) ein Planetengetriebe
ist, ein Sonnenrad (13) des Planetengetriebemechanismus in Eingriff ist, um in eine
Richtung entgegengesetzt zu der einer Ausgangswelle des kontinuierlichen variablen
Getriebes (3) zu drehen, ein Träger (15) des Planetengetriebemechanismus fähig ist,
wahlweise in Eingriff oder von dem Eingriff gelöst zu sein, um in eine Richtung entgegengesetzt
zu der der Ausgangswelle des Verbrennungsmotors (1) zu drehen, das Sonnenrad (13)
und ein Hohlrad (14) fähig sind, wahlweise eingegliedert oder außer Eingriff zu sein,
der erste Schaltmodus durch Lösens des Trägers (15) und der Ausgangswelle des Verbrennungsmotors
(1) und Eingliedern des Sonnenrads (13) und des Hohlrads (14) gesetzt wird, und der
zweite Schaltmodus, der einen Bereich des Getriebeverhältnisses, der von dem des ersten
Schaltmodus verschiedenen ist, durch außer Eingriff Bringen des Sonnenrads (13) und
des Hohlrads (14) und Eingreifen des Trägers (15) und der Ausgangswelle des Verbrennungsmotors
(1) gesetzt wird.
15. Steuerverfahren nach Anspruch 13 und 14, dadurch gekennzeichnet, dass
der Schaltmodus umgeschaltet wird, wenn das Eingangs-/Ausgangsumdrehungsverhältnis,
das durch die Eingangs-/Ausgangsumdrehungsverhältnisberechnungseinrichtung (S5) berechnet
wird, innerhalb des Bereichs des Gesamtgetriebeverhältnisses ist, der durch einen
anderen Schaltmodus als einen gegenwärtigen Schaltmodus definiert ist.
16. Steuerverfahren nach einem der Ansprüche 13 bis 15,
dadurch gekennzeichnet, dass
der Schaltmodus umgeschaltet wird und das berechnete Eingangs-/Ausgangsumdrehungsverhältnis
gesetzt wird, nachdem ein gegenwärtiges Eingangs-/Ausgangsumdrehungsverhältnis auf
einen vorgegebenen Wert variiert wird, der zum Umschalten des Schaltmodus geeignet
ist.
17. Steuerverfahren nach Anspruch 17,
dadurch gekennzeichnet, dass
das Gesamtgetriebeverhältnis gesteuert wird, um eine am effizienteste Drehzahl des
Verbrennungsmotors (1) zu erreichen, und der Verbrennungsmotor (1) gesteuert wird,
so dass der Verbrennungsmotor (1) ein Drehmoment gemäß dem erforderlichen Ausgang
generiert.
18. Steuerverfahren nach Anspruch 16, ferner
gekennzeichnet durch die Schritte:
Abschätzen einer Zeit für das Eingangs-/Ausgangsumdrehungsverhältnis (γ), um den vorgegebenen
Wert zu erreichen, wenn das Eingangs-/Ausgangsumdrehungsverhältnis (γ) variiert wird,
während ein Umschalten des Schaltmodus involviert ist; und
Bestimmen einer Zeitgebung eines Startens einer Schaltmodusumschaltsteuerung auf der
Grundlage der abgeschätzten Zeit.
19. Steuerverfahren nach Anspruch 18,
dadurch gekennzeichnet, dass
die Zeit für das Eingangs-/Ausgangsumdrehungsverhältnis (γ), um den vorgegebenen Wert
zu erreichen, auf der Grundlage einer Variation je Zeiteinheit in dem Eingangs-/Ausgangsumdrehungsverhältnis
(γ) abgeschätzt wird.
20. Steuerverfahren nach Anspruch 19,
dadurch gekennzeichnet, dass
der Schaltmodus durch Eingliedern oder außer Eingriff Bringen des Sonnenrads (13)
und des Hohlrads (14) und durch Eingreifen oder Lösens des Trägers (15) und der Ausgangswelle
des Verbrennungsmotors (1) auf der Grundlage der Zeit, die durch die Abschätzeinrichtung
abgeschätzt wird, umgeschaltet wird.
21. Steuerverfahren nach Anspruch 18 oder 20,
dadurch gekennzeichnet, dass
die Steuerung ferner einen Drehzahlsensor zum Messen einer Eingangsdrehzahl und einer
Ausgangsdrehzahl des kontinuierlichen variablen Getriebes (3); und
einen Zeitmesser zum Messen eines Zeitintervalls zum Erfassen der Drehzahlen hat,
und wenn die nachstehende Gleichung erfüllt ist, bestimmt wird, dass die Umschaltsteuerung
zu starten ist:
wobei T das Zeitintervall ist, γ
n-1 und γ
n das Eingangs-/Ausgangsumdrehungsverhältnis sind, das aus den Drehzahlen berechnet
wird, die bei dem Zeitintervall T erfasst werden, γ
min das vorgegebene Eingangs-/Ausgangsumdrehungsverhältnis ist, und Tst eine Zeit ist,
die erforderlich ist, um den Schaltmodus umzuschalten.
22. Steuerverfahren nach Anspruch 16, ferner gekennzeichnet durch den Schritt
Ändern der Zeitgebung im Wesentlichen eines Startens der Schaltmodusumschaltsteuerung
auf der Grundlage einer Dauer des vorgegebenen Eingangs-/Ausgangsumdrehungsverhältnisses
in einem Prozess eines Variierens des Eingangs-/Ausgangsumdrehungsverhältnisses, während
ein Umschalten des Schaltmodus involviert ist.
23. Steuerverfahren nach Anspruch 22,
dadurch gekennzeichnet, dass
die Dauer des vorgegebenen Eingangs-/Ausgangsumdrehungsverhältnisses auf der Grundlage
einer vergangenen Betriebsaufzeichnung der Dauer korrigiert wird.
24. Steuerverfahren nach entweder Anspruch 22 oder 23,
dadurch gekennzeichnet, dass
die Steuerung ferner einen Drehzahlsensor zum Messen einer Eingangsdrehzahl und einer
Ausgangsdrehzahl des kontinuierlich variablen Getriebes (3); und
einen Zeitmesser zum Messen eines Zeitintervalls zum Erfassen der Drehzahlen hat,
und, wenn die nachstehende Gleichung erfüllt ist, bestimmt wird, dass die Umschaltsteuerung
zu starten ist:
wobei T ein Zeitintervall ist, γ
n-1 und γ
n das Eingangs-/Ausgangsdrehverhältnis sind, das aus den Drehzahlen berechnet wird,
die bei dem Zeitintervall T erfasst werden, γ
min das vorgegebene Eingangs-/Ausgangsumdrehungsverhältnis ist, Tst eine Zeit ist, die
erforderlich ist, um den Schaltmodus umzuschalten, und ΔT ein Korrekturwert auf der
Grundlage der vergangenen Betriebsaufzeichnung der Dauer ist.
1. Contrôleur configuré pour commander un moteur à combustion interne (1) et une boîte
de vitesses (3, 4) couplée à une sortie du moteur à combustion interne (1) et incluant
une boîte de vitesses à variation continue (3) et un mécanisme de levier de vitesses
(4), et le contrôleur est en mesure de régler sélectivement un premier mode de changement
de vitesse (L) dans lequel un rapport de vitesse global (Γ) augmente avec une augmentation
d'un rapport de rotation d'entrée/sortie (γ), à savoir, un rapport entre une vitesse
de rotation d'entrée et une vitesse de rotation de sortie de la boîte de vitesses
à variation continue (3), et un deuxième mode de changement de vitesse (H) dans lequel
le rapport de vitesse global (Γ) diminue avec une augmentation du rapport de rotation
d'entrée/sortie (γ), comprenant :
un moyen de commande du moteur à combustion interne (6) destiné à obtenir un couple
cible sur la base d'une donnée de commande incluant une sortie requise, et à commander
une charge du moteur à combustion interne (1) de manière à atteindre le couple cible
; caractérisé par
un moyen de calcul du rapport de rotation d'entrée/sortie (S5) destiné à obtenir une
vitesse de rotation d'entrée cible sur la base de la donnée de commande et d'une plage
de consommation de carburant optimale et à calculer le rapport de rotation d'entrée/sortie
de la boîte de vitesses à variation continue (3) de manière à atteindre la vitesse
de rotation d'entrée cible ; et
un moyen de détermination du mode de changement de vitesse (S5) destiné à déterminer
un mode de changement de vitesse devant être réglé à partir du premier mode de changement
de vitesse (L) et du deuxième mode de changement de vitesse (H) sur la base du rapport
de rotation d'entrée/sortie (γ) calculé par le moyen de calcul du rapport de rotation
d'entrée/sortie (S5), dans lequel ledit deuxième mode de changement de vitesse (H)
est défini si ledit rapport de vitesse global (Γ) est inférieur à un rapport de rotation
d'entrée/sortie (γ) le plus petit possible de seulement ladite boîte de vitesses à
variation continue (3).
2. Contrôleur selon la revendication 1,
caractérisé en ce que
un arbre d'entrée de la boîte de vitesses à variation continue (3) est couplé à un
arbre de sortie du moteur à combustion interne (1), le mécanisme de levier de vitesses
(4) est un mécanisme d'engrenage planétaire, un planétaire (13) du mécanisme d'engrenage
planétaire est en prise de manière à tourner dans une direction contraire à celle
d'un arbre de sortie de la boîte de vitesses à variation continue (3), un plateau
porteur (15) d'un mécanisme d'engrenage planétaire est en mesure d'être sélectivement
en prise ou libéré de manière à tourner dans une direction contraire à celle de l'arbre
de sortie du moteur à combustion interne (1), le planétaire (13) et une couronne (14)
sont en mesure d'être sélectivement intégrés ou désengagés, le premier mode de changement
de vitesse est réglé en libérant le plateau porteur (15) et l'arbre de sortie du moteur
à combustion interne (1) et en intégrant le planétaire (13) et la couronne (14), et
le deuxième mode de changement de vitesse qui définit une plage du rapport de vitesse
qui est différente de celle du premier mode de changement de vitesse est réglé en
désengageant le planétaire (13) et la couronne (14) et en mettant en prise le plateau
porteur (15) et l'arbre de sortie du moteur à combustion interne (1).
3. Contrôleur selon la revendication 1 ou 2,
caractérisé en ce que
le mode de changement de vitesse est changé lorsque le rapport de rotation d'entrée/sortie
calculé par le moyen de calcul du rapport de rotation d'entrée/sortie (S5) se situe
dans la plage du rapport de vitesse global défini par un mode de changement de vitesse
autre qu'un mode de changement de vitesse en cours.
4. Contrôleur selon l'une quelconque des revendications 1 à 3, caractérisé en ce qu'il
comprend en outre un moyen de commande de la boîte de vitesses à variation continue
(S12) destiné à changer le mode de changement de vitesse et régler le rapport de rotation
d'entrée/sortie calculé (γ) après avoir modifié un rapport de rotation d'entrée/sortie
en cours jusqu'à une valeur prédéterminée appropriée pour changer le mode de changement
de vitesse.
5. Contrôleur selon la revendication 4,
caractérisé en ce que
le moyen de commande de la boîte de vitesses à variation continue (S12) commande le
rapport de vitesse global (Γ) de manière à atteindre une vitesse de rotation la plus
efficace du moteur à combustion interne (1), et le moyen de commande du moteur à combustion
interne (1) commande le moteur à combustion interne (1) de manière à ce que le moteur
à combustion interne (1) génère un couple en fonction de la sortie requise.
6. Contrôleur selon la revendication 4,
caractérisé en ce qu'il comprend en outre :
un moyen d'estimation (S101, S111) destiné à estimer un temps pour que le rapport
de rotation d'entrée/sortie (γ) atteigne la valeur prédéterminée lors de la modification
du rapport de rotation d'entrée/sortie (γ) tout en impliquant le changement du mode
de changement de vitesse ; et
un moyen de commande de changement du mode de changement de vitesse destiné à déterminer
un moment de démarrage de la commande de changement du mode de changement de vitesse
sur la base du temps estimé par le moyen d'estimation (S101, S111).
7. Contrôleur selon la revendication 6,
caractérisé en ce que
le moyen d'estimation (S101, S111) estime le temps pour que le rapport de rotation
d'entrée/sortie (γ) atteigne la valeur prédéterminée sur la base d'une variation par
unité de temps dans le rapport de rotation d'entrée/sortie (γ).
8. Contrôleur selon la revendication 7,
caractérisé en ce que
le moyen de commande de changement du mode de changement de vitesse intègre ou désengage
le planétaire (13) et la couronne (14), et met en prise ou libère le plateau porteur
(15) et l'arbre de sortie du moteur à combustion interne (1) sur la base du temps
estimé par le moyen d'estimation.
9. Contrôleur selon l'une quelconque des revendications 6 et 8,
caractérisé en ce que
le contrôleur comprend en outre un capteur de rotation destiné à détecter une vitesse
de rotation d'entrée et une vitesse de rotation de sortie de la boîte de vitesses
à variation continue (3) ; et
un minuteur destiné à mesurer un intervalle de temps pour détecter les vitesses de
rotation, et quand l'équation suivante est remplie, il est déterminé que la commande
de changement doit être démarrée :
où T est l'intervalle de temps,
γ
n-1 et γ
n sont le rapport de rotation d'entrée/sortie calculé à partir des vitesses de rotation
détectées à l'intervalle de temps T, γ
min est le rapport de rotation d'entrée/sortie prédéterminé, et Tst est le temps requis
pour changer le mode de changement de vitesse.
10. Contrôleur selon la revendication 4, caractérisé en ce qu'il comprend en outre un moyen de commande de changement du mode de changement de vitesse
destiné à changer le moment de démarrage sensible de la commande de changement du
mode de changement de vitesse sur la base d'une durée du rapport de rotation d'entrée/sortie
prédéterminé dans un processus de modification du rapport de rotation d'entrée/sortie
tout en impliquant le changement du mode de changement de vitesse.
11. Contrôleur selon la revendication 10, caractérisé en ce que
la durée du rapport de rotation d'entrée/sortie prédéterminé est corrigée sur la base
d'un enregistrement opérationnel précédent de la durée.
12. Contrôleur selon l'une ou l'autre des revendications 10 et 11,
caractérisé en ce que
le contrôleur comprend en outre un capteur de rotation destiné à mesurer une vitesse
de rotation d'entrée et une vitesse de rotation de sortie de la boîte de vitesses
à variation continue (3) ; et
un minuteur destiné à mesurer un intervalle de temps pour détecter les vitesses de
rotation, et lorsque l'équation suivante est remplie, il est déterminé que la commande
de changement doit être démarrée :
où T est l'intervalle de temps,
γ
n-1 et γ
n sont le rapport de rotation d'entrée/sortie calculé à partir des vitesses de rotation
détectées à l'intervalle de temps T, γ
min est le rapport de rotation d'entrée/sortie prédéterminé, Tst est le temps requis
pour changer le mode de changement de vitesse, et ΔT est une valeur de correction
basée sur l'enregistrement opérationnel précédent de la durée.
13. Procédé de commande d'un moteur à combustion interne (1) et d'une boîte de vitesses
(3, 4), dans lequel ladite boîte de vitesses couplée à une sortie du moteur à combustion
interne (1) inclut la boîte de vitesses à variation continue (3) et un mécanisme de
levier de vitesses (4), le procédé étant en mesure de régler sélectivement un premier
mode de changement de vitesse (L) dans lequel un rapport de vitesse global (Γ) augmente
avec une augmentation du rapport de rotation d'entrée/sortie (γ), à savoir, un rapport
entre une vitesse de rotation d'entrée et une vitesse de rotation de sortie de la
boîte de vitesses à variation continue (3), et un deuxième mode de changement de vitesse
(H) dans lequel le rapport de vitesse global (Γ) diminue avec une augmentation du
rapport de rotation d'entrée/sortie (γ), comprenant l'étape consistant à :
obtenir un couple cible sur la base d'une donnée de commande incluant une sortie requise,
et commander une charge du moteur à combustion interne (1) de manière à atteindre
le couple cible ;
caractérisé par les étapes consistant à :
obtenir une vitesse de rotation d'entrée cible sur la base de la donnée de commande
et d'une plage de consommation de carburant optimale et calculer le rapport de rotation
d'entrée/sortie (γ) de la boîte de vitesses à variation continue (3) de manière à
atteindre la vitesse de rotation d'entrée cible ; et
déterminer un mode de changement de vitesse devant être réglé à partir du premier
mode de changement de vitesse (L) et du deuxième mode de changement de vitesse (H)
sur la base du rapport de rotation d'entrée/sortie calculé, dans lequel ledit deuxième
mode de changement de vitesse (H) est réglé si ledit rapport de vitesse global (Γ)
est inférieur à un rapport de rotation d'entrée/sortie (γ) le plus petit possible
de seulement ladite boîte de vitesses à variation continue (3).
14. Procédé de commande selon la revendication 13,
caractérisé en ce que
un arbre d'entrée de la boîte de vitesses à variation continue (3) est couplé à un
arbre de sortie du moteur à combustion interne (1), le mécanisme de levier de vitesses
(4) est un mécanisme d'engrenage planétaire, un planétaire (13) du mécanisme d'engrenage
planétaire est en prise de manière à tourner dans une direction contraire à celle
d'un arbre de sortie de la boîte de vitesses à variation continue (3), un plateau
porteur (15) du mécanisme d'engrenage planétaire est en mesure d'être sélectivement
en prise ou libéré de manière à tourner dans une direction contraire à celle de l'arbre
de sortie du moteur à combustion interne (1), le planétaire (13) et une couronne (14)
sont en mesure d'être sélectivement intégrés ou désengagés, le premier mode de changement
de vitesse est réglé en libérant le plateau porteur (15) et l'arbre de sortie du moteur
à combustion interne (1) et en intégrant le planétaire (13) et la couronne (14), et
le deuxième mode de changement de vitesse qui définit une plage du rapport de vitesse
qui est différente de celle du premier mode de changement de vitesse est déterminé
en désengageant le planétaire (13) et la couronne (14) et en mettant en prise le plateau
porteur (15) et l'arbre de sortie du moteur à combustion interne (1).
15. Procédé de commande selon la revendication 13 et 14,
caractérisé en ce que
le mode de changement de vitesse est changé lorsque le rapport de rotation d'entrée/sortie
calculé par le moyen de calcul du rapport de rotation d'entrée/sortie (S5) se situe
dans la plage du rapport de vitesse global défini par un mode de changement de vitesse
autre qu'un mode de changement de vitesse en cours.
16. Procédé de commande selon l'une quelconque des revendications 13 à 15, caractérisé en ce que
le mode de changement de vitesse est changé, et le rapport de rotation d'entrée/sortie
calculé est réglé après qu'un rapport de rotation d'entrée/sortie en cours a été modifié
jusqu'à une valeur prédéterminée appropriée pour changer le mode de changement de
vitesse.
17. Procédé de commande selon la revendication 17,
caractérisé en ce que
le rapport de vitesse global est commandé de manière à atteindre une vitesse de rotation
la plus efficace du moteur à combustion interne (1), et le moteur à combustion interne
(1) est commandé de manière à ce que le moteur à combustion interne (1) génère un
couple en fonction de la sortie requise.
18. Procédé de commande selon la revendication 16,
caractérisé en ce qu'il comprend en outre les étapes consistant à :
estimer un temps pour que le rapport de rotation d'entrée/sortie (γ) atteigne la valeur
prédéterminée lors de la modification du rapport de rotation d'entrée/sortie (γ) tout
en impliquant le changement du mode de changement de vitesse ; et
déterminer un instant de démarrage de la commande de changement du mode de changement
de vitesse sur la base du temps estimé.
19. Procédé de commande selon la revendication 18,
caractérisé en ce que
le temps pour que le rapport de rotation d'entrée/sortie (γ) atteigne la valeur prédéterminée
est estimé sur la base d'une variation par unité de temps dans le rapport de rotation
d'entrée/sortie (γ).
20. Procédé de commande selon la revendication 19,
caractérisé en ce que
le mode de changement de vitesse est changé en intégrant ou désengageant le planétaire
(13) et la couronne (14), et en mettant en prise ou libérant le plateau porteur (15)
et l'arbre de sortie du moteur à combustion interne (1) sur la base du temps estimé
par le moyen d'estimation.
21. Procédé de commande selon l'une quelconque des revendications 18 et 20,
caractérisé en ce que
le contrôleur comprend en outre un capteur de rotation destiné à mesurer une vitesse
de rotation d'entrée et une vitesse de rotation de sortie de la boîte de vitesses
à variation continue (3) ; et
un minuteur destiné à mesurer un intervalle de temps pour détecter les vitesses de
rotation, et quand l'équation suivante est remplie, il est déterminée que la commande
de changement doit être démarrée :
où T est l'intervalle de temps,
γ
n-1 et γ
n sont le rapport de rotation d'entrée/sortie calculé à partir des vitesses de rotation
détectées à l'intervalle de temps T, γ
min est le rapport de rotation d'entrée/sortie prédéterminé, Tst est le temps requis
pour changer le mode de changement de vitesse.
22. Procédé de commande selon la revendication 16, caractérisé en ce qu'il comprend en outre l'étape consistant à changer le moment de démarrage sensible
de la commande de changement du mode de changement de vitesse sur la base d'une durée
du rapport de rotation d'entrée/sortie prédéterminé dans un processus de modification
du rapport de rotation d'entrée/sortie tout en impliquant le changement du mode de
changement de vitesse.
23. Procédé de commande selon la revendication 22,
caractérisé en ce que
la durée du rapport de rotation d'entrée/sortie prédéterminé est corrigée sur la base
d'un enregistrement opérationnel précédent de la durée.
24. Procédé de commande selon l'une ou l'autre des revendications 22 et 23,
caractérisé en ce que
le contrôleur comprend en outre un capteur de rotation destiné à mesurer une vitesse
de rotation d'entrée et une vitesse de rotation de sortie de la boîte de vitesses
à variation continue (3) ; et
un minuteur destiné à mesurer un intervalle de temps pour détecter les vitesses de
rotation, et lorsque l'équation suivante est remplie, il est déterminé que la commande
de changement doit être démarrée :
où T est l'intervalle de temps,
γ
n-1 et γ
n sont le rapport de rotation d'entrée/sortie calculé à partir des vitesses de rotation
détectées à l'intervalle de temps T, γ
min est le rapport de rotation d'entrée/sortie prédéterminé, Tst est le temps requis
pour changer le mode de changement de vitesse, et ΔT est une valeur de correction
basée sur l'enregistrement opérationnel précédent de la durée.